Direct actuated valve control hydraulic pump and motor

ABSTRACT

A control device for controlling a displacement pump or motor, comprising a variable mechanical device, comprising at least two masks, wherein the at least two masks are phased relative to each other and a shaft, and are configured to create a desired cam profile. A cam follower is operatively connected to at least one of an intake valve an output valve of a displacement pump or motor.

CROSS-REFERENCE TO RELATED APPLICATIONS

The present application is related to and claims the priority benefit ofU.S. Provisional Patent Application Ser. No. 62/408,131, filed Oct. 14,2016, the contents of which are hereby incorporated by reference intheir entirety into this disclosure.

GOVERNMENT RIGHTS

This invention was made with government support under EEC0540834 awardedby the National Science Foundation. The government has certain rights inthe invention.

TECHNICAL FIELD

The present disclosure generally relates to pumps and motors, and inparticular to use of a variable mechanical device to directly actuatevalves of a hydraulic pump/motor unit to achieve flow diverting and flowlimiting displacement control.

BACKGROUND

This section introduces aspects that may help facilitate a betterunderstanding of the disclosure. Accordingly, these statements are to beread in this light and are not to be understood as admissions about whatis or is not prior art.

A large problem facing hydraulic fluid pumps is low overall efficiencyespecially at low displacements. Digital displacement pumping techniqueshave been developed to mitigate such inefficiencies. Current digitalpump/motors use solenoid poppet valves or a solenoid valve latchingtechnique. The most common units (swashplate axial units) use valveplates. Current digital units also require significant and expensivecontrollers for the solenoids used in the other techniques. Therefore,improvements are needed in the field.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a diagram of a flow diverting system according to oneembodiment.

FIG. 2A shows a state diagram for valve cycling in a first state.

FIG. 2B shows a state diagram for valve cycling in a second state.

FIG. 3 shows a diagram of a pump being mechanically actuated using a camfollower.

FIG. 4A shows a diagram of a variable cam having two masks in a firstrotation state.

FIG. 4B shows a diagram of a variable cam having two masks in a secondrotation state.

FIG. 5A shows a first portion of a simulation model of a half maskingcam pump according to one embodiment.

FIG. 5B shows a second portion of a simulation model of a half maskingcam pump according to one embodiment.

FIG. 5C shows a third portion of a simulation model of a half maskingcam pump according to one embodiment.

FIG. 6 is a graph showing transition length and compression angle of thecam profile.

FIG. 7A is a graph showing an efficiency contour plot at 20%displacement in a first example.

FIG. 7B is a graph showing an efficiency contour plot at 40%displacement in a first example.

FIG. 7C is a graph showing an efficiency contour plot at 60%displacement in a first example.

FIG. 7D is a graph showing an efficiency contour plot at 80%displacement in a first example.

FIG. 7E is a graph showing an efficiency contour plot at 100%displacement in a first example.

FIG. 8A is a graph showing an efficiency contour plot at 20%displacement in a second example.

FIG. 8B is a graph showing an efficiency contour plot at 40%displacement in a second example.

FIG. 8C is a graph showing an efficiency contour plot at 60%displacement in a second example.

FIG. 8D is a graph showing an efficiency contour plot at 80%displacement in a second example.

FIG. 8E is a graph showing an efficiency contour plot at 100%displacement in a second example.

FIG. 9 is a graph showing pressure angles for an example cam accordingto one embodiment.

FIG. 10 is a graph showing operating conditions on the distribution oflosses at 20% displacement.

FIG. 11 shows an example on/off cam actuator and pump arrangementaccording to one embodiment.

FIG. 12 shows an example on/off check pumping cam actuator and pumparrangement according to one embodiment.

FIG. 13 shows an example cam FC bi-directional cam actuator and pumparrangement according to one embodiment.

FIG. 14 shows an example fixed cam FD bi-directional actuator and pumparrangement according to one embodiment.

FIG. 15 shows an example directional on/off cam actuator and pumparrangement according to one embodiment.

FIG. 16 shows an example bi-directional cam actuator and pumparrangement according to one embodiment.

DETAILED DESCRIPTION

For the purposes of promoting an understanding of the principles of thepresent disclosure, reference will now be made to the embodimentsillustrated in the drawings, and specific language will be used todescribe the same. It will nevertheless be understood that no limitationof the scope of this disclosure is thereby intended.

In response to the unmet need, the design disclosed herein addressessuch issues by using state of the art operating strategies and directlyactuating the valves using a variable, mechanical device which reduceslubricating gaps and consequently leakage and therefore increasesefficiency compared with conventional devices. Additionally, in manycases the hydraulic fluid is not needed for lubrication at all whichallows for many different fluids to be used within the unit. Thistechnique is applicable to many different types of devices including butnot limited to radial piston, inline piston, and wobble plate pistonunits. While any variable mechanical actuation could be used, thecurrent design utilizes two half masks which are phased relative to oneanother and the shaft in order to create the desired cam profile. Thisis then followed by the valve poppet to open and close the valveallowing for significant and consistent actuation stroke with minimalcontrol input and power losses.

Efficiency is the main attraction of the herein disclosed methods anddevices. Traditional unit efficiency has good efficiency at highdisplacement which drops off quickly as displacement is reduced. Usingthe flow diverting and flow limiting variable displacement strategiescombined with this technology, efficiency will more closely scale withdisplacement and will lead to greater efficiency especially indisplacement controlled systems. Also, as one embodiment of a functionof the flow limiting/diverting strategies, the piston chamber is onlypressurized when doing work further reducing losses. Additionally,actuating the valves mechanically compared to electronically can supplylarger flow areas (less metering losses) and greater consistency inactuation. The spring return on the follower also reclaims some energyfor actuation. Compared to standard pump/motors (namely swashplate axialunits) leakage losses are greatly reduced as hydraulic fluid is notnecessarily used for lubrication of moving parts. Port plates removed infavor of positively sealing valves. This also allows the hereindescribed methods and devices to be used with non-standard hydraulicfluids such as but not limited to water. Reduces the amount of controlsneeded. Additionally, the herein disclosed methods and devices can beimplemented on most fixed displacement, piston type units with minimalexpense as long as the piston chambers are stationary.

According to one embodiment, a Partial Flow-Diverting strategy isutilized to provide variable displacement. As the name implies, withthis technique excess flow is diverted back to the low pressure portrather than pumping it to the high pressure side; varying the amount offlow which is diverted back allows us to achieve variable displacement.The flow diverting strategy is pictorially illustrated in FIG. 1. As thepiston 100 is moving down starting at top dead center (TDC), valve 01opens allowing the fluid to flow (top right of FIG. 1) and completelyfill the displacement chamber (bottom right of FIG. 1). Valve 101 wouldstill be held open as the piston moves upwards (bottom left of FIG. 1)towards the top dead center (TDC); holding valve 101 open would divertthe flow back to the low pressure port. The displacement of the unit isdefined by the piston position at which valve 101 closes during theupward motion of the piston. When the right amount of fluid is availablein the displacement chamber, valve 101 closes and the fluid ispressurized; then valve 102 opens (upper left of FIG. 1) and the highpressure fluid is pumped into the high pressure line.

According to one embodiment, the repetitive nature of the valveactuation is exploited to allow for mechanical actuation. A stateanalysis was performed on the valves and several configurations from asimple pump to full four-quadrant pump/motors were deemed viable. FIGS.2A and 2B show examples of the state diagrams for the valve cycling.Each diagram represents the state the valve must be in, in this exampleopen or closed, at the corresponding rotation angle.

FIG. 3 shows one example embodiment of a pump 300 with a low pressureintake valve 302 being mechanically actuated using a cam follower 304and a high pressure output valve implemented as a check valve 306. Inthe illustrated configuration, the speed of the valve actuation will beproportional to the speed of the shaft of the pump 300 and the highactuation forces of the cam allow for valves with larger orifice areasto be used. In order to achieve variable displacement of the pump 300the cam profile needs to be varied. To accomplish this, two 50% high camprofiles, or masks, are provided. This allows the effective cam profileto be varied from 100% to 50% high by phasing one mask relative to theother. FIG. 4 illustrates one example of a variable cam having twohalves or masks 402 and 404, with each mask having a 50% high profile.The two halves 402 and 404 can be rotated or phased with respect to oneanother. In FIG. 4A, the two halves are shown in phase with each other,thereby providing a 50% overall high profile. In FIG. 4B, they are shownphased to provide an overall high profile of approximately 85%. The camfollower 304 is implemented as a roller follower to reduce the frictionand size of the cams. The roller follower is connected to the valve 302to achieve opening or closing of the valve as the cam rotates.

In order to determine the most efficient configuration for the halfmasking cam, a simulation of the illustrated pump was created inMatlab/Simscape. The kinematics and pumping chambers used in this modelare based on those created for simulation of an electronicallycontrolled digital pump motor. A single piston version of themechanically actuated Simulink model is shown in FIGS. 5A, 5B, and 5C. Asingle piston is shown for clarity; the simulation was run with threepistons.

The model of FIGS. 5A, 5B and 5C allows for complete manipulation of thecam profiles and allows these to be tested under any operating parameterwithin the scope of pump operation. The main variables to be manipulatedare the transition length, compression angle, transition type, pressuredifference, speed, and displacement.

Transition type is the path that is followed to increase or decrease thecam profile from the low state to the high state. While many differenttransition types may be used, harmonic is used in the simulation as itwas determined to have the lowest peak acceleration and jerk. Othertransition types, such as cycloidal and parabolic may also be used.

Transition length is the amount of degrees the transition type takes totransition from the low to high state. An ideal valve would have atransition length of zero resulting in no addition metering as the valveis opening; thus efficiency is inversely proportional to the amount oftransition degrees. Transition length affects the acceleration of thecam follower as well as the pressure angle of the cam. A rule of thumbfor industry is the pressure angle should not exceed 30° in high speedapplications to prevent seizing. The pressure angle can be calculatedusing equation (1) below where e is eccentricity, x is the camdisplacement profile, and r_(b) is the base radius of the cam.

$\begin{matrix}{\alpha_{p} = {\tan^{- 1}\left( \frac{{x^{\prime}(\theta)} - e}{{x(\theta)} + \sqrt{r_{b}^{2} - e^{2}}} \right)}} & (1)\end{matrix}$

Compression angle is the amount of degrees of rotation the pistoncylinder needed to decompress the fluid as the piston chambertransitions from expulsion to the high pressure line to intake from thelow pressure line at TDC. FIG. 6 shows transition length (θ_(t)) andcompression angle (θ_(c)) on the cam profile.

The correct compression angle is necessary to reclaim the energy inputinto the fluid and maximize efficiency. For an ideal valve thecompression angle would be calculated using the following equations.While this is still useful for estimation, the transition type andlength will affect the ideal compression angle on a real machine.Equation (22) shows the relationship between the change in volume andthe change in pressure.

$\begin{matrix}{{dP} = {K\frac{dV}{V}}} & (2)\end{matrix}$

where V is the effective volume of the piston chamber which isequivalent to the dead volume and the piston area, A_(p), multiplied bythe piston stroke ((3). It should be noted that at TDC, x=1 removing thesecond part of the equation. The change in volume, dV, is equivalent tothe change in piston stroke multiplied by the piston area (4).

V=V _(dead) +A _(p)(l−x)  (3)

dV=Δx A _(p)  (4)

Combining these three equations to determine the compression angle forTDC, (5 is derived.

$\begin{matrix}{{\Delta \; x} = \frac{{dP}\mspace{14mu} V_{dead}}{K\mspace{14mu} A_{p}}} & (5)\end{matrix}$

Spring constant was calculated using (6 where m_(f) is the mass of thefollower and x_(pc) is the spring pre-compression distance.

$\begin{matrix}{k = {\max \left( {- \frac{{x^{''}(\theta)}*m_{f}}{{x(\theta)} + x_{pc}}} \right)}} & (6)\end{matrix}$

In one example, the following system parameters simulated.

-   -   Valve max opening area=100 mm²    -   Poppet stroke=4 mm    -   Piston cylinder radial gap=16 μm    -   Piston Area=314 mm²    -   Piston Stroke=30 mm    -   Transition Length=6-12°    -   Compression Angle=12-20°    -   Speed=500, 1000, 1500, and 2000 rpm    -   Pressure differential=69 and 138 bar    -   Displacement=20%, 40%, 60%, 80%, 100%    -   Number of Pistons=3.

The results of the simulation demonstrate an increase in overallefficiency. When the compression angle was optimized for the operatingrange of 69-138 bar, an average efficiency of 89.72% was reached at an18° compression angle. The same method used for optimization for thisoperating range could be used with any desired parameters. FIGS. 7A-7Fand 8A-8F show efficiency contour plots for varying displacements at aspeed of 2000 rpm and differing pressures. By comparing these graphs,the following trends with respect to the mechanically actuated pump maybe observed:

-   -   Efficiency decreases as the transition length increases.    -   Increase in pressure increases efficiency    -   Increase in pressure increases the ideal compression angle.    -   There is a distinct efficiency maximum at the ideal compression        angle.

Ideally, the transition length would minimized to maximize efficiency,however, this is practically not feasible. In order to stay below the30° pressure angle requirement while keeping the cam size relativelysmall, the following values were selected in one example:

-   -   Cam base diameter=140 mm    -   Roller diameter=26 mm    -   Eccentricity=20 mm    -   Spring pre-compression=2 mm    -   Follower mass=0.1 kg    -   Transition length=10 degrees

FIG. 9 shows the pressure angles for the cam selected for thissimulation. Notice that the maximum value is 28.47 degrees for therising section of the cam, while falling the pressure angle does reach43.82 degrees. Eccentricity was used to shift the pressure angle in thisway because during the falling section, there is less force acting in adirection that will prevent the cam from seizing.

Using the above parameters, FIG. 10 shows the effect of differentoperating conditions on the distribution of losses when operating at 20%displacement. Notice that viscous friction losses are the losses in thepiston chamber due to friction and are relatively constant for this pumpregardless of operating conditions. For these operating conditions,valve losses remain lower than the frictional losses of the pump thoughspeed is a significant factor.

Referring to FIGS. 11, 12, 13, 14, 15, and 16, further exampleembodiments of mechanical cam actuators for use with the pumps describedabove are shown. It shall be appreciated that these configurations areonly example configurations and are not intended to be limiting on thetypes of configurations claimed using the herein disclosed methods anddevices.

Those skilled in the art will recognize that numerous modifications canbe made to the specific implementations described above. Theimplementations should not be limited to the particular limitationsdescribed. Other implementations may be possible.

1. A method for achieving flow diverting and flow limiting variabledisplacement control of a displacement pump or motor, comprisingutilizing a variable mechanical device to provide a control signal foran intake and output valve operatively connected to said displacementpump or motor.
 2. The method of claim 1, wherein the variable mechanicaldevice comprises a rotating cam connected to a rotating shaft and a camfollower operatively connected to at least one of said intake and outputvalve.
 3. The method of claim 1, wherein the cam comprises at least twomasks, wherein the two masks are phased relative to each other and ashaft, and are configured to create a desired cam profile forcontrolling at least one of said intake and output valve.
 4. A controldevice for controlling a displacement pump or motor, comprising: avariable mechanical device, comprising at least two rotating cam masks,wherein the at least two masks are phased relative to each other and ashaft, and are configured to create a desired cam profile; and a camfollower operatively connected to at least one of an intake valve and anoutput valve of a displacement pump or motor.
 5. The variable mechanicaldevice of claim 3, wherein the at least two masks are half masks.